The power output of many multi-stage steam turbine systems is controlled by throttling the main flow of steam from a steam generator in order to reduce the pressure of steam at the high pressure turbine inlet. Steam turbines which utilize this throttling method are often referred to as full arc turbines because all steam inlet nozzle chambers are active at all load conditions. Full arc turbines are usually designed to accept exact steam conditions at a rated load in order to maximize efficiency. By admitting steam through all of the inlet nozzles, the pressure ratio across the inlet stage, e.g., the first control stage, in a full arc turbine remains essentially constant irrespective of the steam inlet pressure. As a result, the mechanical efficiency of power generation across the control stage may be optimized. However, as power is decreased in a full arc turbine, there is an overall decline in efficiency, i.e., the ideal efficiency of the steam work cycle between the steam generator and the turbine output, because throttling reduces the energy available for performing work. Generally, the overall turbine efficiency, i.e., the actual efficiency is a product of the ideal and the mechanical efficiency of the turbine.
More efficient control of turbine output than is achievable by the throttling method has been realized by the technique of dividing steam which enters the turbine inlet into isolated and individually controllable arcs of admission. In this method, known as partial-arc admission, the number of active first stage nozzles is varied in response to load changes. Partial arc admission turbines have been favored over full arc turbines because a relatively high ideal efficiency is attainable by sequentially admitting steam through individual nozzle chambers with a minimum of throttling, rather than by throttling the entire arc of admission. The benefits of this higher ideal efficiency are generally more advantageous than the optimum mechanical efficiency achievable across the control stage of full arc turbine designs. Overall, multi-stage steam turbine systems which use partial-arc admission to vary power output operate with a higher actual efficiency than systems which throttle steam across a full arc of admission. However, partial-arc admission systems in the past have been known to have certain disadvantages which limit the efficiency of work output across the control stage. Some of these limitations are due to unavoidable mechanical constraints, such as, for example, an unavoidable amount of windage and turbulence which occurs as rotating blades pass nozzle blade groups which are not admitting steam.
Furthermore, in partial-arc admission systems the pressure drop (and therefore the pressure ratio) across the nozzle blade groups varies as steam is sequentially admitted through a greater number of valve chambers, the largest pressure drop occurring at the minimum valve point (fewest possible number of governor or control valves open) and the smallest pressure drop occurring at full admission. The thermodynamic efficiency, which is inversely proportional to the pressure differential across the control stage, is lowest at the minimum valve point and highest at full admission. Thus, the control stage efficiency for partial-arc turbines as well as full arc turbines decreases when power output drops below the rated load. However, given the variable pressure drops across the nozzles of a partial-arc turbine, it is believed that certain design features commonly found in partial-arc admission systems can be improved upon in order to increase the overall efficiency of a turbine. Because the control stage is an impulse stage wherein most of the pressure drop occurs across the stationary nozzles, a one percent improvement in nozzle efficiency will have four times the effect on control stage efficiency as a one percent improvement in the efficiency of the rotating blades. Turbine designs which provide even modest improvements in the performance of the control stage nozzles will significantly improve the actual efficiency of partial-arc turbines. At their rated loads, even a 0.25 percent increase in the actual efficiency of a partial-arc turbine can result in very large energy savings.
Sliding or variable throttle pressure operation of partial-arc turbines also results in improved turbine efficiency and additionally reduces low cycle fatigue. The usual procedure is to initiate sliding pressure operation on a partial-arc admission turbine at flows below the value corresponding to the point where half the control valves are wide open and half are fully closed, i.e., 50% first stage admission on a turbine in which the maximum admission is practically 100%. If sliding pressure is initiated at a higher flow (larger value of first stage admission), there is a loss in performance. However, in a turbine having eight valves, sliding from 75% admission eliminates a considerable portion of the valve loop (valve throttling) on the sixth valve which would occur with constant throttle pressure operation. A similar situation occurs when sliding from 62.5% admission: a considerable portion of the valve loop of the fifth valve is eliminated. Elimination of such valve loops improves the turbine heat rate and its efficiency.
FIG. 1 illustrates the effect of sliding pressure control in a partial-arc steam turbine having eight control valves. The abscissa represents values of steam flow while the ordinate values are heat rate. Line 10 represents constant pressure with throttling control while line 12 represents sliding pressure on a full arc admission turbine. Line 14 represents constant pressure with sequential valve control (partial-arc admission) and dotted lines 16, 18, 20 and 22 represent the valve loops. The valve loops result from gradual throttling of each of a sequence of control or governor valves. Sliding pressure operation from 75% admission is indicated by line 24. Note that much of the valve loop 20 is eliminated by sliding pressure along line 24 but that heat rate (the reciprocal of efficiency) increases disproportionately below the 62.5% admission point. Line 26, showing sliding pressure from the 62.5% admission point, provides some improvement but does not affect valve loops 16, 18 and 20. Similarly, sliding from 50% admission, line 28, helps at the low end but does not affect valve loops 16-22. Each of these valve loops represent higher heat rates and reduced efficiency from the ideal curve represented by line 14.
FIGS. 2, 3 and 4 illustrate the operation of an exemplary steam turbine using one prior art control. FIG. 2 shows the locus of full valve points, line 30, with constant pressure operation at 2535 psia. The valve points are at 50%, 75%, 87.5% and 100% admission with the valve loops identified by the lines 32, 34 and 36. Sliding pressure is indicated by lines 38, 40 and 42. Starting at 100% admission, about 806 MW for the exemplary turbine system, load is initially reduced by keeping all eight control valves wide open and sliding throttle pressure by controlling the steam producing boiler. When the throttle pressure, line 38, reaches the intersection point with the valve loop 32, the throttle pressure is increased to 2535 psia while closing the eighth control valve. The control valve would continue to close as load is further reduced while maintaining the 2535 psia throttle pressure until this valve is completely closed at which point the turbine is operating at 87.5% admission. To further reduce load, valve position is again held constant, seven valves fully open, and throttle pressure is again reduced until the throttle pressure corresponds to the intersection of the sliding pressure line 40 and the valve loop 34 for the seventh valve. To reduce load below this point, the pressure is increased to 2535 psia and the seventh valve is progressively closed (riding down the valve loop) until it is completely closed. The admission is now 75%. To reduce load still further, the pressure is again reduced with six valves wide open and two fully closed until the throttle pressure line 42 reaches the intersection with the valve loop 36 where the fifth and sixth valves move simultaneously with constant throttle pressure operation. Then the operation of raising throttle pressure and closing of the valves is repeated for any number of valves desired. The variation in throttle pressure is illustrated in FIG. 3. The sloped portions 44 of line 46 relates to the sliding pressure regime with constant valve position. The vertical portions 48 relate to the termination of sliding pressure with no valve throttling and the uppermost point relates to operation at full pressure with valve throttling. The horizontal portions 50 relate to the riding down of the valve loop while reducing load at constant pressure. FIG. 4 shows the improvement in heat rate as a function of load. The line 52 illustrates the difference between valve loop performance at constant pressure and the performance with variable pressure between valve points.
The performance improvements shown in FIGS. 2 and 4 are based on the assumption that the boiler feed pump discharge is reduced as the throttle pressure is reduced. If it is not reduced proportionally, the improvement is reduced since the energy required to maintain discharge pressure remains high. In the prior art system, a signal is sent to the feed pump-feed pump drive system to reduce pressure. In reality, however, the feed pump is followed by a pressure regulator in order to eliminate the need for constant adjustment of pump speed and the occurrence of control instability and hunting because of small variations in inlet water pressure to the boiler, resulting from perturbations in flow demand. The regulator, then, does more or less throttling which changes pump discharge pressure and therefore the flow that the pump will deliver. The pump speed is held constant over a desired range of travel of the regulator valve. When the valve travel gets outside these limits, the pump speed is adjusted to move the valve to some de sired mean position. As a consequence, the pump discharge pressure does not equal the minimum allowable value (throttle pressure plus system head losses) and so the performance improvement is not as large as shown by FIGS. 2 and 4. In addition, in order to achieve quicker load response, the regulator valve is usually operated with some pressure drop so that if there is a sudden increase in load demand, the valve can open quickly and increase flow. The response of the pump and its drive is slower than the response of the regulator valve.